The internal combustion engine, i.e. the engine, and the associated ancillaries are a vibratable system, the vibration behavior of which can be influenced. The most relevant vibratable systems or components under shock and force excitation are the crankcase, cylinder block, cylinder head, valve drive and crank drive, which is also the subject of the present invention. These components are exposed to the mass and gas forces.
The temporally changing rotary forces, which are introduced into the crankshaft via the connecting rods mounted pivotingly at the individual crank journals, excite the crankshaft into rotational vibrations. These rotational vibrations lead both to noise from structure-borne sound emission and noise from structure-borne sound transmission into the bodywork and into the internal combustion engine, wherein vibrations can also occur which have a disadvantageous influence on driving comfort, for example vibrations of the steering wheel in the passenger compartment. When the crankshaft is excited in its inherent frequency range, high rotational vibration amplitudes can occur, which can even lead to fatigue fracture. This shows that the vibrations are important not only in connection with noise design, but also with regard to component strength.
The rotational vibrations of the crankshaft are transmitted undesirably to the camshaft via the timing drive or camshaft drive, wherein the camshaft is itself also a vibratable system and can excite further systems, in particular the valve drive, to vibration. Vibrations can also be transmitted to other ancillaries via traction means driven by the crankshaft. Also the vibrations of the crankshaft are transmitted to the drive train, via which they can be transmitted as far as the vehicle tires.
The rotary force development at a crankshaft throw of a four-stroke internal combustion engine is periodic, wherein the periods extend over two revolutions of the crankshaft. Normally the rotary force development is broken down into its harmonic elements by Fourier analysis, to allow conclusions on the excitation of rotary vibrations. The actual rotary force development consists of a constant rotary force and a plurality of harmonically changing rotary forces, which have different rotary force amplitudes and frequencies or vibration counts. The ratio of the vibration count ni of each harmonic to rotation speed n of the crankshaft or engine is known as the order i of the harmonic.
Because of the high dynamic load on the crankshaft from the mass and gas forces, the aim during design of the internal combustion engine is to achieve as extensive, i.e., as optimized a mass compensation as possible. The term “mass compensation” covers all measures which outwardly compensate or reduce the effect of the mass forces. To this extent, mass compensation also comprises compensation of the moments provoked by the mass forces.
A mass compensation can take place in individual cases by a targeted matching of the crankshaft throws and the number and arrangement of the cylinders.
A six-cylinder in-line engine can be fully balanced in this way. The six cylinders are combined in pairs such that they run mechanically in parallel as cylinder pairs. So the first and sixth cylinders, the second and fifth cylinders, and the third and fourth cylinders are combined into cylinder pairs, wherein the crankshaft journals or throws for the three cylinder pairs are arranged on the crankshaft each offset by 120° CA. Running mechanically in parallel means that both pistons of the two cylinders running mechanically in parallel are at top dead center (TDC) and bottom dead center (BDC) at the same ° CA (degree crank angle).
In a three-cylinder in-line engine, the mass forces of the first order and the mass forces of the second order can also be fully compensated by selection of a suitable crankshaft throw, but not the moments which are provoked by the mass forces.
Complete mass compensation is not always achievable, so further measures must be taken, for example the arrangement of counter weights on the crankshaft, and/or equipping the internal combustion engine with at least one balancer shaft.
The starting point of all measures is the consideration that the crankshaft is loaded by the temporally changing rotary forces, composed of the gas forces and mass forces of the crank drive. The mass of the crank drive, i.e., the individual masses of the connecting rod, piston, piston pin, piston rings and the crankshaft itself, can be divided into an oscillating substitute mass and a rotating substitute mass. The external effect of the mass force of the rotating substitute mass can easily be compensated by counter weights arranged on the crankshaft.
Compensation is more difficult in the case of the mass force provoked by the oscillating substitute mass, since this is composed roughly of a mass force of the first order and a mass force of the second order, wherein forces of higher orders are negligible.
The mass forces of each order can be almost compensated by the arrangement of two contra-rotating shafts fitted with corresponding weights, known as balancer shafts. The shafts for compensating for mass forces of the first order run at the engine rotation speed, and the shafts for balancing the mass forces of the second order run at double the engine rotation speed. This method of mass compensation is very cost-intensive, complex as well as having a high weight, and requires a great deal of space. Within the context of compensating for the mass forces of the first order, the crankshaft can simultaneously serve as a balancer shaft, i.e., it can constitute one of the two balancer shafts, so that at least the weight and space required for mass compensation is reduced.
Even when the mass forces are completely compensated, mass moments arise since the mass forces of the individual cylinders act in the cylinder center planes. These mass moments can in individual cases be compensated by at least one balancer shaft equipped with weights. This further increases the space required, the costs, and the weight for the total mass compensation.
In a three-cylinder in-line engine, the moments provoked by the mass forces of the first order are compensated, for example, by a single balancer shaft contra-rotating to the crankshaft at the engine rotation speed, at the ends of which shaft two compensation weights serving as imbalance are arranged offset, i.e., twisted, by 180°.
Alternatively, the moments provoked by the mass forces of the first order in a three-cylinder in-line engine can also be compensated by two contra-rotating compensation weights serving as imbalance, wherein a first compensation weight runs in the same direction as the crankshaft and the second compensation weight runs in the opposite direction to the crankshaft. The essential difference from the mass compensation described above is that the two compensation weights serving as imbalance rotate in opposition to each other. Consequently the two compensation weights are not arranged on the same carrier, for example, a shaft, but on different carriers which give them a rotary motion in different directions of rotation. The carrier for the first compensation weight can, for example be the crankshaft itself or a body connected with the crankshaft, for example a flywheel. The second compensation weight requires a carrier rotating in the opposite direction to the crankshaft, which can itself also be driven by the crankshaft.
As explained in detail above, many concepts for compensating for mass forces and/or mass moments require the arrangement of compensation weights serving as imbalance on the crankshaft.
The external effect of the mass force of the rotating substitute mass for example can be compensated completely by compensation weights arranged on the crankshaft. Here, the at least one compensation weight serving as imbalance is arranged on the crankshaft on the side opposite at least one crankshaft throw, for the purpose of mass compensation.
Also, concepts for compensating for moments provoked by mass forces of the first order, for example of a three-cylinder in-line engine, can use compensation weights serving as imbalance which are arranged on the crankshaft.
The arrangement of compensation weights on the crankshaft in the manner described above can give rise to constructional problems. When the piston passes through bottom dead center, a compensation weight serving as imbalance, and arranged on the side opposite the crankshaft throw belonging to the piston, lies immediately below the piston, i.e., on the side facing the piston. Contact between the piston and the compensation weight as the crankshaft rotates must be prevented. Normally compensation weights have an arcuate form on the outward-facing side, wherein the outside of the compensation weight running in the peripheral direction has a substantially constant distance R from the longitudinal axis of the crankshaft.
The constricted spatial conditions in the crankcase however often make the arrangement of a compensation mass difficult, or require the arrangement of a comparatively large compensation mass, since the effective lever, namely the distance of the center of gravity of the imbalance from the longitudinal axis of the crankshaft, is or must be selected comparatively small, i.e., short, because of the small construction space available.
Where applicable, because of the constricted spatial conditions, mass compensation must be provided outside the crankcase. The provision of one or where applicable several balancer shafts outside the crankcase not only increases the space required in the engine bay of a vehicle and the costs, but also the fuel consumption. The increased fuel consumption is caused firstly by the additional weight of the compensation unit. Secondly, the compensation unit with its rotating shafts and other moving components contributes not insignificantly to the friction power of the internal combustion engine, or increases this friction power. The latter is particularly relevant because of the fact that the compensation unit is always and continuously in operation as soon as the internal combustion engine is started and operated.
With regard to fuel consumption however, the compensation masses and the entire weight of the mass compensation must in principle be as low as possible.